Diagnostic system and method for a cooling system

ABSTRACT

A diagnostic system and method for a cooling system includes producing a control signal for adjusting a valve position of a regulator valve as a demand for cooling, as well as providing an ambient temperature measurement, coil inlet temperature measurement, and coil exit temperature measurement. The valve position and at least one of the ambient temperature measurement, coil inlet temperature measurement, and coil exit temperature measurement are monitored and compared with at least one predetermined fault value indicative of a fault condition to diagnose the system. An alert module responsive to the diagnostic module issues an alert signal when at least one of the ambient temperature measurement, coil inlet temperature measurement, and coil exit temperature measurement bears a predetermined relationship to the fault value. When monitoring and comparing, an error value percentage indicative of a percentage of sampled error within an accepted offset range for a defined period of time is determined.

CROSS-REFERENCE TO RELATED APPLICATIONS

This is a division of U.S. Ser. No. 10/147,782, filed May 16, 2002 nowU.S. Pat. No. 6,499,305, which is a division of U.S. Ser. No.09/886,592, filed Jun. 21, 2001, now U.S. Pat. No. 6,467,280; which is adivision of U.S. Ser. No. 09/524,364, filed Mar. 14, 2000, now U.S. Pat.No. 6,408,635; which is a division of U.S. Ser. No. 08/939,779, filedSep. 29, 1997, now U.S. Pat. No. 6,047,557; which is acontinuation-in-part of U.S. Ser. No. 08/486,118, filed Jun. 7, 1995,now U.S. Pat. No.5,741,120, each of which is incorporated herein byreference in its entirety.

BACKGROUND AND SUMMARY OF THE INVENTION

The present invention relates generally to cooling systems, compressorcontrol systems and refrigerant regulating valve control systems. Moreparticularly, the invention relates to a refrigeration system employinga pulse width modulated compressor or evaporator stepper regulatorcontrolled by a variable duty cycle signal derived from a load sensor.Preferably an adaptive controller generates the variable duty cyclesignal. The compressor has two mechanical elements separated by a seal,and these mechanical elements are cyclically movable relative to oneanother to develop fluid pressure. The compressor includes a mechanismto selectively break the seal in response to the control signal, therebymodulating the capacity of the system.

The refrigeration system can be deployed as a distributed system inrefrigeration cases and the like. The preferred arrangement allows thecompressor and condenser subsystems to be disposed in or mounted on therefrigeration case, thereby greatly reducing the length of refrigerantconduit and refrigerant required.

Conventionally, refrigeration systems for supermarket refrigerationcases have employed air-cooled or water-cooled condensers fed by a rackof compressors. The compressors are coupled in parallel so that they maybe switched on and off in stages to adjust the system cooling capacityto the demands of the load. Commonly, the condensers are locatedoutside, on the roof, or in a machine room adjacent the shopping areawhere the refrigeration cases are located.

Within each refrigeration case is an evaporator fed by lines from thecondensers through which the expanded refrigerant circulates to cool thecase. Conventionally, a closed-loop control system regulates refrigerantflow through the evaporator to maintain the desired case temperature.Proportional-integral-derivative (PID) closed loop control systems arepopular for this purpose, with temperature sensors and/or pressuresensors providing the sensed condition inputs.

It is common practice within supermarkets to use separate systems tosupply different individual cooling temperature ranges: low temperature(for frozen foods, ice cream, nominally −25F.); medium (for meat, dairyproducts, nominally +20F.); high (for floral, produce, nominally +35 to+40F.). The separate low, medium and high temperature systems are eachoptimized to their respective temperature ranges. Normally, each willemploy its own rack of compressors and its own set of refrigerantconduits to and from the compressors and condensers.

The conventional arrangement, described above, is very costly toconstruct and maintain. Much of the cost is associated with the longrefrigerant conduit runs. Not only are long conduit runs expensive interms of hardware and installation costs, but the quantity ofrefrigerant required to fill the conduits is also a significant factor.The longer the conduit run, the more refrigerant required. Adding to thecost are environmental factors. Eventually fittings leak, allowing therefrigerant to escape to atmosphere. Invariably, long conduit runsinvolve more pipefitting joints that may potentially leak. When a leakdoes occur, the longer the conduit run, the more refrigerant lost.

There is considerable interest today in environmentally friendlyrefrigeration systems. Shortening the conduit run is seen as one way toachieve a more environmentally friendly system. To achieve this, newcondenser/compressor configurations and new control systems will need tobe engineered.

Re-engineering condenser/compressor configurations for moreenvironmentally friendly systems is not a simple task, because systemefficiency should not be sacrificed. Generally, the conventionalroof-mounted condenser system, supplied by condensers, benefits fromeconomies of scale and is quite efficient. These systems serve as thebenchmark against which more environmentally friendly systems of thefuture will need to be measured.

To appreciate why re-engineering an environmentally yet efficient systemhas proven so difficult, consider these thermodynamic issues. Thetypical refrigeration case operates in a very unpredictable environment.From a design standpoint, the thermal mass being cooled is rarelyconstant. Within the supermarket environment, the temperature andhumidity may vary widely at different times of day and over differentseasons throughout the year. The product load (items in therefrigeration case) can also change unpredictably. Customers removingproduct and store clerks replenishing product rarely synchronize.Outside the supermarket environment, the outdoor air temperature andhumidity may also vary quite widely between day and night and/or betweensummer and winter. The capacity of the system must be designed for theharshest conditions (when the condenser environment is the hottest).Thus systems may experience excess capacity in less harsh conditions,such as in the cool evenings or during the winter.

Periodic defrosting also introduces thermal fluctuations into thesystem. Unlike thermal fluctuations due to environmental conditions, thethermal fluctuations induced by the defrost cycle are cause by thecontrol system itself and not by the surrounding environment.

In a similar fashion, the control system for handling multiplerefrigeration cases can induce thermal fluctuations that are quitedifficult to predict. If all cases within a multi-case system aresuddenly turned on at once—to meet their respective cooling demands—thecooling capacity must rapidly be ramped up to maximum. Likewise, if allcases are suddenly switched off, the cooling capacity should be rampeddown accordingly. However, given that individual refrigeration cases mayoperate independently of one another, the instantaneous demand forcooling capacity will tend to vary widely and unpredictably.

These are all problems that have made the engineering of environmentallyfriendly systems more difficult. Adding to these difficulties are userengineering/ergonomic problems. The present day PID controller can bedifficult to adapt to distributed refrigeration systems. Experiencedcontrols engineers know that a well-tuned PID controller can involve adegree of artistry in selecting the proper control constants used in thePID algorithm. In a large refrigeration system of the conventionalarchitecture (non-distributed) the size of the system justifies having acontrols engineer visit the site (perhaps repeatedly) to fine tune thecontrol constant parameters.

This may not be practical for distributed systems in which thecomponents are individually of a much smaller scale and far morenumerous. By way of comparison, a conventional system might employ onecontroller for an entire multi-case, store-wide system. A distributedsystem for the same store might involve a controller for each case oradjacent group of cases within the store. Distributed systems need to bedesigned to minimize end user involvement. It would therefore bedesirable if the controller were able to auto configure. Currentlycontrol systems lack this capability.

The present invention provides a distributed refrigeration system inwhich the condenser is disposed on the refrigeration case and servicedby a special pulse width modulated compressor that may be also disposedwithin the case. If desired, the condenser and compressor can be coupledto service a group of adjacent refrigeration cases, each case having itsown evaporator. The pulse width modulated compressor employs twomechanical elements, such as scroll members, that move rotationallyrelative to one another to develop fluid pressure for pumping therefrigerant. The compressor includes a mechanism that will selectivelybreak the seal between the two mechanical elements, thereby altering thefluid pressure developed by the compressor while allowing the mechanicalelements to maintain substantially constant relative movement with oneanother. The compressor can be pulse width modulated by making andbreaking the fluid seal without the need to start and stop the electricmotor driving the mechanical elements.

The pulse width modulated compressor is driven by a control system thatsupplies a variable duty cycle control signal based on measured systemload. The controller may also regulate the frequency (or cycle time) ofthe control signal to minimize pressure fluctuations in the refrigerantsystem. The on time is thus equal to the duty cycle multiplied by thecycle time, where the cycle time is the inverse of the frequency.

The refrigeration system of the invention has a number of advantages.Because the instantaneous capacity of the system is easily regulated byvariable duty cycle control, an oversized compressor can be used toachieve faster temperature pull down at startup and after defrost,without causing short cycling as conventional compressor systems would.Another benefit of variable duty cycle control is that the system canrespond quickly to sudden changes in condenser temperature or casetemperature set point. The controller adjusts capacity in response todisturbances without producing unstable oscillations and withoutsignificant overshoot. Also, the ability to match instantaneous capacityto the demand allows the system to operate at higher evaporatortemperatures. (Deep drops in temperature experienced by conventionalsystems at overcapacity are avoided.)

Operating at higher evaporator temperatures reduces the defrost energyrequired because the system develops frost more slowly at highertemperatures. Also, the time between defrosts can be lengthened by apercentage proportional to the accumulated runtime as dictated by theactual variable duty cycle control signal. For example, a sixty percentduty cycle would increase a standard three-hour time between defrosts tofive hours (3/0.60=5).

The pulse width modulated operation of the system yields improved oilreturn. The refrigerant flow pulsates between high capacity and lowcapacity (e.g. 100% and 0%), creating more turbulence which breaks downthe oil boundary layer in the heat exchangers.

Another benefit of the variable duty cycle control system is its abilityto operate with a variety of expansion devices, including the simpleorifice, the thermal expansion valve (TXV) and the electronic expansionvalve. A signal derived from the expansion device controller can be fedto the compressor controller of the invention. This signal allows thevariable duty cycle control signal and/or its frequency to be adjustedto match the instantaneous operating conditions of the expansion device.A similar approach may be used to operate variable speed fans in aircooled condenser systems. In such case the controller of the inventionmay provide a signal to control fan speed based on the current operatingduty cycle of the compressor.

Yet another benefit of the invention is its ability to detect when thesystem is low on refrigerant charge, an important environmental concern.Low refrigerant charge can indicate the presence of leaks in the system.Low charge may be detected by observing the change in error betweenactual temperature and set point temperature as the system duty cycle ismodulated. The control system may be configured to detect when themodulation in duty cycle does not have the desired effect on temperaturemaintenance. This can be due to a loss of refrigerant charge, a stuckthermal expansion valve or other malfunctions.

For a more complete understanding of the invention, its objects andadvantages, refer to the following specification and to the accompanyingdrawings.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a system block diagram of a prior art refrigeration systemconfiguration;

FIG. 2 is a block diagram of a refrigeration system in accordance withthe present invention;

FIG. 3 is a cross-sectional view of an embodiment of the pulse widthmodulated compressor, shown in the loaded state;

FIG. 4 is a cross-sectional view of the compressor of FIG. 3, shown inthe unloaded state;

FIG. 5 is another embodiment of a refrigeration or cooling system inaccordance with the present invention;

FIG. 6 is a block diagram of the controller;

FIG. 7 is a block diagram showing how the controller may be used tomodulate an evaporator stepper regulator;

FIG. 8 is a block diagram of the signal conditioning module of thecontroller of FIG. 6;

FIG. 9 is a block diagram of the control module of the controller ofFIG. 6;

FIG. 10 is a state diagram depicting the operating states of thecontroller;

FIG. 11 is a flowchart diagram illustrating the presently preferred PIcontrol algorithm;

FIG. 12 is a waveform diagram illustrating the variable duty cyclesignal produced by the controller and illustrating the operation at aconstant frequency;

FIG. 13 is a waveform diagram of the variable duty cycle signal,illustrating variable frequency operation;

FIG. 14 is a series of graphs comparing temperature and pressuredynamics of system employing the invention with a system of conventionaldesign;

FIG. 15 is a block diagram illustrating the adaptive tuning module ofthe invention;

FIG. 16a is a flowchart diagram illustrating the presently preferredoperation of the adaptive tuning module, specifically with respect tothe decision whether to start tuning;

FIG. 16b is a flowchart diagram illustrating the presently preferredprocess performed by the adaptive tuning module in the integration mode;

FIG. 16c is a flowchart diagram illustrating the operation of theadaptive tuning module in the calculation mode;

FIG. 17 is a state diagram illustrating the operative states of theadaptive tuning module;

FIG. 18 is a block diagram illustrating the fuzzy logic block of theadaptive tuning loop;

FIG. 19 is a membership function diagram for the fuzzy logic block ofFIG. 18;

FIG. 20 is a truth table relating to the membership function of FIG. 19as used by the fuzzy logic block of FIG. 18;

FIG. 21 is an output membership function diagram for the fuzzy logicblock of FIG. 18; and

FIG. 22 is a schematic illustrating exemplary sensor locations forcontrol-related and diagnostic-related functions of the presentinvention.

DESCRIPTION OF THE PREFERRED EMBODIMENTS

FIG. 1 illustrates how a conventional supermarket refrigeration systemis configured. As previously discussed, it is conventional practice toplace the compressors 30 and the condenser 32 in a location remote fromthe refrigeration cases 34. In this illustration, the compressors 30 areconfigured in a parallel bank located on the roof 36 of the building.The bank of compressors supply a large condenser 32, which may be aircooled or water cooled. The condenser supplies liquid refrigerant to areceiver 38. The receiver 38, in turn, supplies the individualrefrigeration cases, which are connected in parallel, as illustrated. Inmost implementations a liquid line solenoid valve 40 is used to regulateflow to the associated evaporator 42. The refrigerant is supplied to theevaporator through a suitable expansion device such as expansion valve44. The expansion valve 44 provides a restricted orifice that causes theliquid refrigerant to atomize into liquid droplets that are introducedinto the inlet side of the evaporator 42. The evaporator 42, locatedwithin the refrigeration case 34, extracts heat from the case and itscontents by vaporization of the liquid refrigerant droplets into a gas.The compressors 30 extract this gas by suction and compress it back intothe liquid state. The liquid refrigerant is then cooled in the condenser32 and returned to the receiver 38, whereupon the cycle continues.

To match cooling capacity to the load, the compressors 30 may beswitched on and off individually or in groups, as required. In a typicalsupermarket arrangement there may be several independent systems, eachconfigured as shown in FIG. 1, to handle different operating temperatureranges. Note that the liquid line 46 and the suction line 48 may eachneed to be quite lengthy (e.g., up to 150 feet) to span the distancefrom refrigeration case to roof.

FIG. 2 shows a refrigeration case 34 configured according to theprinciples of the present invention. The condenser 32 and compressor 30are both disposed within case 34 or attached thereto. Evaporator 42 andthe associated expansion valve 44 are likewise disposed within case 34.The condenser 32 is provided with a heat removal mechanism 50 by whichheat is transferred to ambient. The heat removal mechanism can be awater jacket connected to suitable plumbing for carrying waste heat to awater cooling tower located on the building roof. Alternatively, theheat removal mechanism can be a forced-air cooling system or a passiveconvection-air cooling system.

The refrigeration system of the invention employs a compressorcontroller 52 that supplies a pulse width modulated control signal online 54 to a solenoid valve 56 on compressor 30. The compressorcontroller adjusts the pulse width of the control signal using analgorithm described below. A suitable load sensor such as temperaturesensor 58 supplies the input signal used by the controller to determinepulse width.

FIGS. 3 and 4 show the details of compressor 30. FIG. 3 shows thecompressor in its loaded state and FIG. 4 shows the compressor in itsunloaded state. The solenoid valve 56 switches the compressor betweenthese two states while the compressor motor remains energized. Oneimportant advantage of this configuration is that the compressor can bepulse width modulated very rapidly between the loaded and unloadedstates without interrupting power to the compressor motor. This pulsewidth modulated cycling exerts less wear on the compressor, because themotor is not subjected to sudden changes in angular momentum.

Referring to FIGS. 3 and 4, there is shown an exemplary compressor 30.Compressor 30 may be used within a hermetic scroll compressor such asgenerally of the type described in assignee's U.S. Pat. No. 5,102,316.

The exemplary compressor 30 includes an outer shell 61 and an orbitingscroll member 64 supported on upper bearing housing 63 and drivinglyconnect to crankshaft 62 via crank pin 65 and drive bushing 60. A secondnon-orbiting scroll member 67 is positioned in meshing engagement withscroll member 64 and axially movably secured to upper bearing housing63. A partition plate 69 is provided adjacent the upper end of shell 61and serves to define a discharge chamber 70 at the upper end thereof.

In operation, as orbiting scroll member 64 orbits with respect to scrollmember 67, suction gas is drawn into shell 61 via suction inlet 71 andthence into compressor 30 through inlet 72 provided in non-orbitingscroll member 67. The intermeshing wraps provided on scroll members 64and 67 define moving fluid pockets which progressively decrease in sizeand move radially inwardly as a result of the orbiting motion of scrollmember 64 thus compressing the suction gas entering via inlet 72. Thecompressed gas is then discharged into discharge chamber 70 viadischarge port 73 provided in scroll member 67 and passage 74.

In order to unload compressor 30, solenoid valve 56 will be actuated inresponse to a signal from control module 87 to interrupt fluidcommunication to increase the pressure within chamber 77 to that of thedischarge gas. The biasing force resulting from this discharge pressurewill overcome the sealing biasing force thereby causing scroll member 67to move axially upwardly away from orbiting scroll member 64. This axialmovement will result in the creation of a leakage path between therespective wrap tips and end plates of scroll members 64 and 67 therebysubstantially eliminating continued compression of the suction gas.

A flexible fluid line 91 extends from the outer end of passage 90 to afitting 92 extending through shell 61 with a second line 93 connectingfitting 92 to solenoid valve 56. Solenoid valve 56 has fluid lines 82and 84 connected to suction line 83 and discharge line 85 and iscontrolled by control module 87 in response to conditions sensed bysensor 88 to effect movement of non-orbiting scroll member 67 betweenthe positions shown in FIGS. 3 and 4.

When compression of the suction gas is to be resumed, solenoid valve 56will be actuated so as to move scroll member 67 into sealing engagementwith scroll member 64.

The refrigeration case embodiment of FIG. 2 may be packaged as aself-contained unit. While that may be a desirable configuration formany applications, the invention is not restricted to stand alone,self-contained refrigeration case configurations. Rather, the inventionlends itself to a variety of different distributed refrigerationsystems. FIG. 5 shows an example of such a distributed system.

Referring to FIG. 5, a single compressor 30 and condenser 32 can serviceseveral distributed refrigeration cases or several distributed coolingunits in a heating and cooling (HVAC) system. In FIG. 5 therefrigeration cases or cooling system housings are shown as dashedboxes, designated 34 a, 34 b, and 34 c. Conveniently, the compressor 30and condenser 32 may be disposed within or attached to one of therefrigeration cases or housings, such as refrigeration case or housing34 a.

Each refrigeration case or housing has its own evaporator and associatedexpansion valve as illustrated at 42 (a, b, c) and 44 (a, b, c). Inaddition, each refrigeration case or housing may have its owntemperature sensor 58 (a, b, c) supplying input information to thecompressor controller 52. Finally, a pressure sensor 60 monitors thepressure of the suction line 48 and supplies this information tocompressor controller 52. The compressor controller supplies a variableduty cycle signal to the solenoid valve 56 as previously described.

The multiple case or multiple cooling unit embodiment of FIG. 5 showshow a single compressor can be pulse width modulated by compressorcontroller 52 to supply the instantaneous demand for cooling. Thetemperature sensors 58 (a, b, c) collectively provide an indication ofthe load on the system, as does pressure sensor 60. The controlleradjusts the pulse width of the control signal to modulate the compressorbetween its high capacity and low capacity states (100%, 0%) to meet theinstantaneous demand for refrigerant.

As an alternate control technique, one or more of the suction linesexiting the evaporator can be equipped with an electrically controlledvalve, such as an evaporator pressure regulator valve 45 c. Valve 45 cis coupled to controller 52, as illustrated. It may be supplied with asuitable control signal, depending on the type of the valve. A steppermotor valve may be used for this purpose, in which case controller 30would supply a suitable signal to increment or decrement the setting ofthe stepper motor to thereby adjust the orifice size of the valve.Alternatively, a pulse width modulated valve could be used, in whichcase it may be controlled with the same variable duty cycle signal assupplied to the compressor 30.

Controller 52 is not limited to solely compressor control applications.The variable duty cycle control signal can also be used to control othertypes of refrigerant flow and pressure control devices, such asrefrigerant regulating valves. FIG. 7 shows such an application, wherethe output of controller 52 supplies control signals to evaporatorstepper regulator 43. This device is a fluid pressure regulator that isadjusted by stepper motor 45. The evaporator stepper regulator (ESR)valve 43 adjusts the suction pressure to thereby adjust the capacity ofthe system.

A block diagram of the presently preferred compressor controller isillustrated in FIG. 6. A description of the various signals and datavalues shown in this and successive figures is summarized in Table Ibelow.

TABLE I Default No. Variable Name Value Description 1 SignalConditioning: Sensor Alarm False Indicates Sensor Reading is not withinexpected range Sensor Mode Min User configuration to indicate ifMin/Max/Avg is performed for all temperature Sensors Sampling Time (Ts)0.5 sec Rate at which Signal condition- ing block is executed ControlType T/P Type if controlled by only Temp. or both Temp. & Pressure 2Control Block: Sensor Alarm False Same as before System Alarm FalseGenerated by Adaptive Block indicative some system problem SSL 0 Steadystate loading % Defrost Status False Whether system is in defrostPull_Down_Time 0 Time taken to pull down after defrost Gain (K) 7 Gainused in PI algorithm Integral Time (Ti) 100 used in PI Control Time (Tc)10 Sec used in PI Control Set Pt. (St) 0° F. used in PI Operating State1 What state the machine is operating at 3 Defrost Control DefrostStatus False If defrost status of the case Defrost Type External If thedefrost is from external timer or Internal clock of controller DefrostInterval 8 hrs Time between defrost Defrost Duration 1 hr DefrostDuration Defrost Termination 50° F. Termination temperature for Temp.defrost

At the heart of the controller is control block module 102. This moduleis responsible for supplying the variable duty cycle control signal onlead 104. Module 102 also supplies the compressor ON/OFF signal on lead106 and an operating state command signal on lead 108. The compressorON/OFF signal drives the contactors that supply operating current to thecompressor motor. The operating state signal indicates what state thestate machine (FIG. 10) is in currently.

The control block module receives inputs from several sources, includingtemperature and pressure readings from the temperature and pressuresensors previously described. These temperature readings are passedthrough signal conditioning module 110, the details of which are shownin the pseudocode Appendix. The control block module also receives adefrost status signal from defrost control module 112. Defrost controlmodule 112 contains logic to determine when defrost is performed. Thepresent embodiment allows defrost to be controlled either by an externallogic signal (supplied through lead 114) or by an internal logic signalgenerated by the defrost control module itself. The choice of whether touse external or internal defrost control logic is user selectablethrough user input 116. The internal defrost control uses user-suppliedparameters supplied through user input 118.

The preferred compressor controller in one form is autoconfigurable. Thecontroller includes an optional adaptive tuning module 120 thatautomatically adjusts the control algorithm parameters (the proportionalconstant K) based on operating conditions of the system. The adaptivetuning module senses the percent loading (on lead 104) and the operatingstate (on lead 108) as well as the measured temperature after signalconditioning (on lead 122). Module 120 supplies the adaptive tuningparameters to control block 102, as illustrated. The current embodimentsupplies proportional constant K on lead 124 and SSL parameter on lead126, indicative of steady-state loading percent. A system alarm signalon lead 126 alerts the control block module when the system is notresponding as expected to changes in the adaptively tuned parameters.The alarm thus signals when there may be a system malfunction or loss ofrefrigerant charge. The alarm can trigger more sophisticated diagnosticroutines, if desired. The compressor controller provides a number ofuser interface points through which user-supplied settings are input.The defrost type (internal/external) input 116 and the internal defrostparameters on input 118 have already been discussed. A user input 128allows the user to specify the temperature set point to the adaptivetuning module 120. The same information is supplied on user input 130 tothe control block module 102. The user can also interact directly withthe control block module in a number of ways. User input 132 allows theuser to switch the compressor on or off during defrost mode. User input134 allows the user to specify the initial controller parameters,including the initial proportional constant K. The proportional constantK may thereafter be modified by the adaptive tuning module 120. Userinput 136 allows the user to specify the pressure differential (dP) thatthe system uses as a set point.

In addition to these user inputs, several user inputs are provided forinteracting with the signal conditioning module 110. User input 138selects the sensor mode of operation for the signal conditioning module.This will be described in more detail below. User input 140 allows theuser to specify the sampling time used by the signal conditioningmodule. User input 142 allows the user to specify whether the controllershall be operated using temperature sensors only (T) or temperature andpressure sensors (T/P).

Referring now to FIG. 8, the signal conditioning module is shown indetail. The inputs (temperature and/or pressure sensors) are showndiagrammatically at 144. These inputs are processed through analog todigital convertor 146 and then supplied to the control type selector148. Temperature readings from the temperature and/or pressure sensorsare taken sequentially and supplied serially through the analog todigital convertor. The control type selector codes or stores the data sothat pressure and temperature values are properly interpreted.

Digital filtering is then applied to the signal at 150 to removespurious fluctuations and noise. Next, the data are checked in module152 to ensure that all readings are within expected sensor range limits.This may be done by converting the digital count data to thecorresponding temperature or pressure values and checking these valuesagainst the pre-stored sensor range limits. If the readings are notwithin sensor range an alarm signal is generated for output on output154.

Next a data manipulation operation is performed at 156 to supply thetemperature and/or pressure data in the form selected by the sensor modeuser input 138. The current embodiment will selectively average the dataor determine the minimum or maximum of the data (Min/Max/Avg). TheMin/Max/Avg mode can be used to calculate the swing in pressuredifferential, or a conditioned temperature value. The average mode canbe used to supply a conditioned temperature value. These are shown asoutputs 158 and 160, respectively.

FIG. 9 shows the control block module in greater detail. The conditionedtemperature or pressure signal is fed to calculation module 162 thatcalculates the error between the actual temperature or pressure and theset point temperature or pressure. Module 162 also calculates the rateof change in those values.

The control block module is designed to update the operating state ofthe system on a periodic basis (every T_(c) seconds, nominally onceevery second). The Find Operating State module 164 performs this updatefunction. The state diagram of FIG. 10 provides the details on how thisis performed. Essentially, the operating state advances, from state tostate, based on whether there is a sensor alarm (SA) present, whetherthere is a defrost status signal (DS) present and what the calculatederror value is. The Find Operating State module 164 supplies theoperating state parameter and the Pull Down Time parameter to thedecision logic module 166.

Referring to FIG. 10, the Find Operating State module 164 advances fromstate to state as follows. Beginning in the initial state 168 the moduleadvances to the normal operating state 170 after initialization. Itremains in that state until certain conditions are met. FIG. 10 shows bylabel arrows what conditions are required to cycle from the normaloperating state 170 to the defrost state 172; to the pull down state174; to the sensor alarm pull down state 176; to the sensor alarmoperating state 178 and to the sensor alarm defrost state 180.

The decision logic module 166 (FIG. 9) determines the duty cycle of thevariable duty cycle signal. This is output on lead 182, designated %Loading. The decision logic module also generates the compressor ON/OFFsignal on lead 184. The actual decision logic will be described below inconnection with FIG. 11. The decision logic module is form ofproportional integral (PI) control that is based on an adaptivelycalculated cycle time T_(cyc). This cycle time is calculated by thecalculation module 186 based on a calculated error value generated bymodule 188. Referring back to FIG. 6, the conditioned pressuredifferential signal on lead 122 (Cond dP) is supplied to the CalculateError module 188 (FIG. 9) along with the pressure differential set pointvalue as supplied through user input 136 (FIG. 6). The differencebetween actual and set point pressure differentials is calculated bymodule 188 and fed to the calculation module 186. The adaptive cycletime T_(cyc) is a function of the pressure differential error and theoperating state as determined by the find operating state module 164according to the following calculation:

T _(cyc(new)) =T _(cyc(old)) +K _(c)* Error  (1)

where:

K_(c): proportional constant; and

Error: (actual-set point) suction pressure swing.

The presently preferred PI control algorithm implemented by the decisionlogic module 166 is illustrated in FIG. 11. The routine begins at step200 by reading the user supplied parameters K, T_(i), T_(c) and S_(t).See FIG. 6 for a description of these user supplied values. The constantK_(p) is calculated as being equal to the initially supplied value K;and the constant K_(i) is calculated as the product of the initiallysupplied constant K and the ratio T_(c)/T_(i).

Next, at step 202 a decision is made whether the absolute value of theerror between set point temperature and conditioned temperature (on lead190, FIG. 9) is greater than 5° F. If so, the constant K_(p) is setequal to zero in step 204. If not, the routine simply proceeds to step206 where a new loading percent value is calculated as described by theequation in step 206 of FIG. 11. If the load percent is greater than 100(step 208), then the load percent is set equal to 100% at step 210. Ifthe load percent is not greater than 100% but is less than 0% (step 212)the load percent is set equal to 0% at step 214. If the load percent isbetween the 0% and 100% limits, the load percent is set equal to the newload percent at step 216.

The variable duty cycle control signal generated by the controller cantake several forms. FIGS. 12 and 13 give two examples. FIG. 12 shows thevariable duty cycle signal in which the duty cycle varies, but thefrequency remains constant. In FIG. 12, note that the cycle time,indicated by hash marks 220, are equally spaced. By comparison, FIG. 13illustrates the variable duty cycle signal wherein the frequency is alsovaried. In FIG. 13, note that the hash marks 220 are not equally spaced.Rather, the waveform exhibits regions of constant frequency, regions ofincreasing frequency and regions of decreasing frequency. The variablefrequency illustrated in FIG. 13 is the result of the adaptivemodulation of the cycle time T_(cyc).

FIG. 14 graphically shows the benefits that the control system of theinvention has in maintaining tighter temperature control and highersuction pressure with improved system efficiency. Note how thetemperature curve 222 of the invention exhibits considerably lessfluctuation than the corresponding temperature curve 224 of aconventional controller. Similarly, note that the pressure curve 226 ofthe invention has a baseline well above that of pressure curve 228 ofthe conventional controller. Also, the peak-to-peak fluctuation inpressure exhibited by the invention (curve 226) is much smaller thanthat of the conventional controller (curve 228).

The controller of the invention operates at a rate that is at least fourtimes faster (typically on the order of at least eight times faster)than the thermal time constant of the load. In the presently preferredembodiment the cycle time of the variable duty cycle signal is abouteight times shorter than the time constant of the load. By way ofnon-limiting example, the cycle time of the variable duty cycle signalmight be on the order of 10 to 15 seconds, whereas the time constant ofthe system being cooled might be on the order of 1 to 3 minutes. Thethermal time constant of a system being cooled is generally dictated byphysical or thermodynamic properties of the system. Although variousmodels can be used to describe the physical or thermodynamic response ofa heating or cooling system, the following analysis will demonstratesthe principle.

Modeling the thermal time constant of the system being cooled.

One can model the temperature change across the evaporator coil of arefrigeration system or heat pump as a first order system, wherein thetemperature change may be modeled according to the following equation:

ΔT=ΔT _(ss)[1−exp(−t/γ)]+ΔT ₀ exp(−τ/γ).

where:

ΔT=air temperature change across coil

ΔT_(ss)=steady state air temperature change across coil

ΔT₀=air temperature change across the coil at time zero

t=time

γ=time constant of coil.

The transient capacity of the unit can be obtained by multiplying theabove equation by the air mass flow rate (m) and specific heat atconstant pressure (C_(p)) and integrating with respect to time.

Generally, it is the removal of the refrigerant from the evaporator thatcontrols the time required to reach steady state operating condition,and thus the steady state temperature change across the condenser coil.If desired, the system can be modeled using two time constants, onebased on the mass of the coil and another based on the time required toget the excess refrigerant from the evaporator into the rest of thesystem. In addition, it may also be desirable to take into account, as afurther time delay, the time lag due to the large physical distancebetween evaporator and condenser coils in some systems.

The thermal response of the evaporator coil may be modeled by thefollowing equation:

=½[(1−e ^(t/γ1))+(1−e ^(t/γ2))]

where:

=temperature change across coil/steady state temperature change acrosscoil

t=time

γ₁=time constant based on mass of coil

γ₂=time constant based on time required to remove excess refrigerantfrom evaporator

In practice, the controller of the invention cycles at a ratesignificantly faster than conventional controllers. This is because theconventional controller cycles on and off in direct response to thecomparison of actual and set-point temperatures (or pressures). In otherwords, the conventional controller cycles on when there is demand forcooling, and cycles off when the error between actual and set-pointtemperature is below a predetermined limit. Thus the on-off cycle of theconventional controller is very highly dependent on the time constant ofthe system being cooled.

In contrast, the controller of the invention cycles on and off at a ratedictated by calculated values that are not directly tied to theinstantaneous relation between actual and set-point temperatures orpressures. Rather, the cycle time is dictated by both the cycle rate andthe duty cycle of the variable duty cycle signal supplied by thecontroller. Notably, the point at which the controller cycles from on tooff in each cycle is not necessarily the point at which the demand forcooling has been met, but rather the point dictated by the duty cycleneeded to meet the demand.

Adaptive Tuning

The controller Geneva described above can be configured to perform aclassic control algorithm, such as a conventionalproportional-integral-derivative (PID) control algorithm. In theconventional configuration the user would typically need to set thecontrol parameters through suitable programming. The controller may alsobe of an adaptive type, described here, to eliminate the need for theuser to determine and program the proper control parameters.

Thus, one important advantage of the adaptive controller is its abilityto perform adaptive tuning. In general, tuning involves selecting theappropriate control parameters so that the closed loop system is stableover a wide range of operating conditions, responds quickly to reducethe effect of disturbance on the control loop and does not causeexcessive wear of mechanical components through continuous cycling.These are often mutually exclusive criteria, and a compromise mustgenerally be made. In FIG. 18 (and also FIG. 6) there are two basiccontrol loops: the refrigeration control loop and the adaptive tuningloop. The refrigeration control loop is administered by control blockmodule 102; the adaptive loop is administered by adaptive tuning module120. Details of the adaptive tuning module 120 are shown in FIGS. 15, 16a-16 c and 17. The presently preferred adaptive tuning module uses afuzzy logic control algorithm that will be described in connection withFIGS. 18-20.

Referring to FIG. 15, the adaptive tuning module performs basicallythree functions. First, it decides whether to perform adaptive tuning.This is handled by module 240. Second, it gathers the needed parametersfor performing adaptive tuning. This is handled by module 242. Third, itcalculates the adaptive gain used by the control loop. This is handledby module 244.

Module 240 bases the decision on whether to start tuning upon twofactors: the current operating state of the system and the control setpoint. The flowchart of FIG. 16a shows the steps involved in thisdecision. Module 242 integrates key parameters needed for thecalculations performed by module 244. Essentially, module 242 inputs thepercent loading, the temperature and pressure values and the set pointtemperature. It outputs the following data: S_ER (the total number ofconditioned temperature and pressure data points that are within 0.5degrees or 1 Psig of the set point value), S_Close (the total number ofpercent loading data points that goes to zero percent during a givensampling interval, e.g. 30 min.), S_Open (the total number of percentloading data points that goes to 100% in the sampling interval) and SSLP(a moving average or rolling average of the percent loading during thesampling interval). Module 242 is responsive to a tuning flag that isset by module 240. Module 242 performs the integration of these keyparameters when signaled to do so by the tuning flag. FIG. 16b shows thesteps involved in performing integration of these key parameters.

Finally, the calculation block takes the data supplied by module 242 andcalculates the adaptive gain using the process illustrated in FIG. 16c.

The adaptive tuning module 120 will cycle through various operatingstates, depending on the state of a timer. FIG. 17 is a state diagramshowing how the presently preferred embodiment will function. Note thatthe sequence transitions from the initialization mode to either theintegration mode or the no tuning mode, depending on whether the tuningflag has been set. Once in the integration mode, the system performsintegration until the timer lapses (nominally 30 minutes), whereupon thecalculation mode is entered. Once the calculations are completed thetimer is reset and the system returns to the initialization mode.

The block diagram of the adaptive scheme is shown in FIG. 18. There aretwo basic loops—The first one is the PID control loop 260 that runsevery “dt” second and the second is the adaptive loop 262 that runsevery “ta” second. When the control system starts, the PID control loop260 uses a default value of gain (K) to calculate the control output.The adaptive loop 262, checks the error e(t) 264 every “ta” seconds 266(preferably less than 0.2* dt seconds). At module 268 if the absolutevalue of error, e(t), is less than desired offset (OS), a counter Er_newis incremented. The Offset (OS) is the acceptable steady-state error(e.g. for temperature control it may be +/1° F.). This checking processcontinues for “tsum” seconds 270 (preferably 200 to 500 times dtseconds). After “tsum” seconds 270, the value Er_new is converted intopercentage (Er_new % 272). The parameter Er_new % 272 indicates thepercentage of sampled e(t) that was within accepted offset (OS) for“tsum” time. In other words, it is a measure of how well the controlvariable was controlled for past “tsum” seconds. A value of 100% means“tight” control and 0% means “poor” control. Whenever Er_new % is 100%,the gain remains substantially unchanged as it indicates tightercontrol. However, if Er_new happens to be between 0 and 100%, adaptivefuzzy-logic algorithm module 274 calculates a new gain (K_new 276) thatis used for next “tsum” seconds by the control algorithm module 278.

In the preferred embodiment, there is one output and two inputs to thefuzzy-logic algorithm module 274. The output is the new gain (K_new)calculated using the input, Er_new %, and a variable, Dir, defined asfollows:

Dir=Sign[(ER−new %−ER_old %)*(K_new−K−old)]  (2)

where:

Sign stands for the sign (+ve, −ve or zero) of the term inside thebracket;

Er_new % is the percentage of e(t) that is within the offset for past“tsum” seconds;

Er_old % is the value of Er_new % in “(tsum−1)” iteration;

K_new is the gain used in “tsum” time; and

K_old is the gain in (tsum−1) time.

For example, suppose the controller starts at 0 seconds with a defaultvalue of K=10 and, ta=1 seconds, tsum=1000 seconds and OS=1. Suppose 600e(t) data out of a possible 1000 data was within the offset. Therefore,after 1000 sec, Er_new %=60 (i.e., 600/1000*100), K_new=10. Er_old % andK_old is set to zero when the adaptive fuzzy-logic algorithm module 274is used the first time. Plugging these numbers in Eq.(2) gives the signof the variable “Dir” as positive. Accordingly, the inputs to theadaptive fuzzy-logic module 274 for the first iteration arerespectively, Er_new %=60 and Dir=+ve.

The next step is to perform fuzzification of these inputs into fuzzyinputs by using membership functions.

Fuzzification:

A membership function is a mapping between the universe of discourse(x-axis) and the grade space (y-axis). The universe of discourse is therange of possible values for the inputs or outputs. For ER_new % it ispreferably from 0 to 100. The value in the grade space typically rangesfrom 0 to 1 and is called a fuzzy input, truth value, or a degree ofmembership. FIG. 19 shows graph 300 which contains the membershipfunctions for the input, Er_new %. Er_new % is divided into threelinguistic variables—LARGE (304), MEDIUM (306) AND SMALL (308). ForEr_new %=60, the fuzzy inputs (or degree of membership function)are—0.25 of LARGE and 0.75 of MEDIUM. The input variable “Dir” is welldefined (+ve, −ve or zero) and thus does not require a membershipfunction in this application. The next step is to create the “TruthTable” or Rule Evaluation.

Rule Evaluation:

Rule evaluation takes the fuzzy inputs from the fuzzification step andthe rules from the knowledge base and calculates fuzzy outputs. FIG. 20shows the rules as truth table. For the first column and first row, therule is:

“IF ER_new % is LARGE AND Dir is NEGATIVE THEN New Gain is NO CHANGE(NC)” (i.e. if the percentage of e(t) data that is within the offset(OS) for last “tsum” seconds is LARGE and the direction (DIR) isNEGATIVE/ZERO then do not change the existing K value (NO CHANGE)).

In the example, because ER_new % has fuzzy inputs LARGE (0.25) ANDMEDIUM (0.75) with POSITIVE Dir, the rules that will be used are:

IF ER_new % is LARGE (0.25) AND Dir is POSITIVE THEN New Gain is NOCHANGE (NC=1)

IF ER_new % is MEDIUM (0.75) AND Dir is POSITIVE THEN New Gain isPOSITIVE SMALL CHANGE (PSC=1.2)

Defuzzification:

Finally, the defuzzification process converts the fuzzy outputs from therule evaluation step into the final output by using Graph 310 of FIG.21. Graph 310, uses the following labels=“NBC” for negative big change;“NSC” for negative small change; “NC” for no change; “PSC” for positivesmall change; and PBC for positive big change. The Center of Gravity orcentroid method is used in the preferred embodiment for defuzzification.The output membership function for change in gain is shown in FIG. 21.

The centroid (the Fuzzy-Logic Output) is calculated as:${Centroid} = {{K\_ new} \cdot \left\lbrack \frac{\sum\limits_{{all} \cdot x}{{\mu (x)} \cdot x}}{\sum\limits_{{all} \cdot x}{\mu (x)}} \right\rbrack}$

where:

(x) is the fuzzy output value for universe of discourse value x.

In our example, the output (K_new) becomes${Output} = {{10 \cdot \left\lbrack \frac{{0.25(1)} + {0.75(1.2)}}{0.25 + 0.75} \right\rbrack} \approx 11.50}$

Once the three steps of fuzzification, rule evaluation, anddefuzzification are finished and the output has been calculated, theprocess is repeated again for new set of Er_new %.

In the above example, after the first 1000 sec, the adaptive algorithmcalculates a new gain of K_new=11.50. This new gain is used for the next1000 sec (i.e. from t=1000 to 2000 sec in real time) by the PID controlloop. At t=1001 sec, counter Er_new is set to zero to perform countingfor the next 1000 seconds. At the end of another 1000 seconds (ie. att=2000 seconds), Er_new % is calculated again.

Suppose this time, Er_new % happens to be 25. This means, by changing Kfrom 10 to 11.5, the control became worse. Therefore, it would be betterto change gain in the other direction, i.e., decrease the gain ratherthan increase. Thus, at t=2000 sec, Er_new %=25, Er_old %=60 (previousvalue of Er_new %), K_new=11.5 and K_(—old=)10 (previous value of K).Applying Eq.(2), a negative “Dir” is obtained. With Er_new % of 25 andDir=Negative, the fuzzy-logic calculation is performed again tocalculate a new gain for the next 1000 seconds. The new value of gain isK_new=7.76 and is used from t=2000 to 3000 seconds by the PID Loop.

Suppose for the third iteration, i.e., from t=2000 to 3000 seconds,Er_new % comes out to be 95% (which represents tighter control).Performing the same fuzzy-logic operation gives the same value of K_new,and the gain remains unchanged until Er_new % again degrades.

Exemplary Applications:

Both pulse width modulated (PWM) Compressors and electronic stepperregulator (ESR) Valves can be used to control evaporatortemperature/pressure or evaporator cooling fluid (air or water)temperature. The former controls by modulating the refrigerant flow andthe latter restricts the suction side to control the flow. Referringback to FIG. 18, the block diagram of the control system for such anactuator working in a refrigeration system 279 is shown. In FIG. 18 oneand preferably up to four temperatures of evaporator cooling fluid orone evaporator suction pressure (generally shown at 282) is sampledevery dt seconds. A sampling time of dt=10 seconds was found to beoptimum for both the applications. After processing by the analog todigital module 284, the sampled signal is then reduced to one number bytaking the average or the minimum or the maximum of the fourtemperatures depending on the system configuration or the userpreference at module 286. Typically, in a single actuator (PWM/ESR)systems where the complete evaporator coil goes into defrost at onetime, averaging of control signal is preferred. In a multipleevaporator-single actuator system where defrost of evaporator coils doesnot occur at the same time, minimum is the preferred mode. The valueobtained after avg/min/max is called conditioned signal. At comparisonmodule 288 this is compared with the desired set point to calculate theerror, e(t).

The control algorithm used in the loop is a Proportion-integral (PI)control technique (PID). The Pi algorithm calculates the valve position(0-100%) in case of ESR or calculates the percentage loading (0 to 100%)in case of PWM compressor. A typical integral reset time, Ti, for boththe actuators is 60 seconds. The gain is tuned adaptively by theadaptive loop. The adaptive algorithm is turned off in the preferredembodiment whenever: the system is in defrost; is going throughpull-down; there a big set point change; sensor failure has beendetected; or any other system failure is detected.

Consequently, the adaptive algorithm is typically used when the systemis working under normal mode. The time “ta” preferably used is about 1seconds and “tsum” is about 1800 seconds (30 minutes).

Diagnostics Related to PWM Compressor/ESR Valves:

Referring to FIG. 22, a discharge cooling fluid temperature sensor 312(Ta), an evaporator coil inlet temperature sensor 314 (Ti) and anevaporator coil exit temperature sensor 316 (To) can provide diagnosticfeatures for the evaporator control using PWM/ESR. The Inlet temperaturesensor 314 can be anywhere in evaporator coil 318. However, thepreferred location is about one third of the total evaporator lengthfrom the evaporator coil distributor 320.

Using these three temperature sensors, system learning can be performedthat can be used for diagnostics. For example, diagnostics can beperformed for ESR/PWM when it is used in a single evaporator along withan expansion valve. In this example, the following variables are trackedevery “tsum” second in the adaptive loop. The variables can beintegrated just after ER_new integration is done in the adaptive loop.

N-Close: Number of times Valve position /PWM loading was 0%.

N-Open: Number of times Valve position/PWM loading was 100%.

MAVP: The moving average of the Valve position /PWM loading for “tsum”seconds.

SSLP: The steady-state Valve position/PWM loading is set equal to MAVPif for the “tsum” duration ER_new % is greater than 50%.

dT: Moving average of the difference between Ta and Ti (Ta-Ti).

SH: Moving average of the difference between To and Ti (To-Ti) in thesaid duration. This is approximately the evaporator superheat.

N_FL: Number of times To was less than Ti during the said duration,i.e., “tsum” seconds. This number will indicate how much the expansionvalve is flooding the evaporator.

In addition, Pull-down time after defrost, tpd, is also learnt. Based onthese variables, the following diagnostics are performed: temperaturesensor failure; degraded expansion valve; degraded ESR valve/PWMCompressor; oversized ESR/PWM; undersized ESR/PWM; and no air flow.

Temperature Sensor Failure:

Failures of temperature sensors are detected by checking whether thetemperature reading falls within the expected range. If PWM/ESR iscontrolled using Ta as the control variable, then when it fails, thecontrol is done as follows. The above said actuator is controlled basedon Ti, or the Ta values are estimated using the learned dT (i.e., add dTto Ti value to estimate Ta). During pull down, the valve/PWM can be setto full-open/load for the learned pull-down time (tpd). If Ti also failsat the same time or is not available, the actuator is opened 100% duringpull down time and then set to steady-state loading percent (SSLP) afterpull-down-time. An alarm is sent to the supervisor upon such acondition.

Degraded Expansion Valve:

If an expansion valve sticks or is off-tuned or is undersized/oversized,the following combinations of the tracked variable can be used todiagnose such problems. N_FL>50% and ER_new %>10% indicate the expansionvalve is stuck open or is off-tuned or may be even oversized and thus isflooding the evaporator coil. An alarm is sent upon such a condition.Moreover, SH>20 and N_FL=0% indicate an off-tuned expansion valve or anundersized valve or the valve is stuck closed.

Degraded ESR Valve/PWM Compressor:

A degraded ESR is one that misses steps or is stuck. A degraded PWMCompressor is one whose solenoid is stuck closed or stuck open. Theseproblems are detected in a configuration where defrost is performed bysetting the ESR/PWM to 0%. The problem is detected as follows.

If ER_new %>50% before defrost and during defrost Ti<32□F and SH>5□F,then the valve is determined to be missing steps. Accordingly, the valveis closed by another 100% and if Ti and SH remain the same then this ishighly indicative that the valve is stuck.

If ER_new %=0 and N_Close is 100% and Ti<32F and SH>5F then PWM/ESR isdetermined to be stuck open. If ER_new %=0 and N_Open is 100% and Ti>32Fand SH>5F then PWM/ESR is determined to be stuck closed.

Over-Sized ESR/PWM:

If N_Close>90% and 30%<ER_new %<100%, then an alarm is sent foroversized valve/PWM Compressor.

Under-Sized ESR/PWM:

If N_Open>90% and ER_new %=0 and SH>5, then an alarm is sent forundersized valve/PWM Compressor.

No Air Flow:

If N_Open=100%, ER_new %=0, SH<5F and Ti<25F and N_FL>50%, then eitherthe air is blocked or the fans are not working properly.

Additionally, these diagnostic strategies can also be applied to anelectronic expansion valve controller.

The embodiments which have been set forth above were for the purpose ofillustration and were not intended to limit the invention. It will beappreciated by those skilled in the art that various changes andmodifications may be made to the embodiments discussed in thisspecification without departing from the spirit and scope of theinvention as defined by the appended claims.

Appendix Pseudocode for performing signal conditioning Repeat thefollowing every Ts Seconds: Read User Inputs: Sampling Time (Ts) ControlType (P or T) Sensor Mode (Avg/Min/Max) Perform Analog to DigitalConversion (ADC) on all (four) Temp. Sensor Channels output data asCounts Digitally Filter Counts Ynew = 0.75 * Yold + 0.25 * Counts outputdata as Filtered Counts Convert Filtered Counts to Deg F. Test if atleast one Sensor is within normal operating limits e.g. within −40 and+90 F. If none are within limit--Set Sensor Alarm to TRUE Else PerformAvg/Min/Max operation based on Sensor Mode If Control Type is NOT a T/PControl Type Then End Signal Conditioning Routine (until next Ts cycle)Else (Control Type is T/P) Do the Following: Perform ADC on PressureSensor Channel output data as Counts Digitally Filter Counts Ynew =0.75 * Yold + 0.25 * Counts output data as Filtered Counts ConvertFiltered Counts to Psig Test if pressure Sensor is within normaloperating limits e.g. within 0 and +200 If not within limit: Set dp = dPSet Pt. Else: Calculate dP = Pmax − Pmin Set Sensor Alarm to ConditionedT/dP End Signal Conditioning Routine (until next Ts cycle)

We claim:
 1. A diagnostic system for a cooling system including a compressor, evaporator and condenser in fluid communication, the diagnostic system comprising: a controller associated with a regulator valve and producing a control signal for adjusting a valve position of said regulator valve as a function of demand for cooling; an ambient temperature sensor associated with the cooling system and operable to provide an ambient temperature measurement to said controller; an evaporator coil inlet temperature sensor associated with the cooling system and operable to provide coil inlet temperature measurement to said controller; an evaporator coil exit temperature sensor associated with the cooling system and operable to provide a coil exit temperature measurement to said controller; a diagnostic module coupled to said controller for monitoring and comparing a valve position and at least one of said ambient temperature measurement, coil inlet temperature measurement and coil exit temperature measurement with at least one predetermined fault value indicative of a fault condition.
 2. The diagnostic system of claim 1, further comprising an alert module responsive to said diagnostic module for issuing an alert signal when said at least one of said ambient temperature measurement, coil inlet temperature measurement and coil exit temperature measurement bears a predetermined relationship to said fault value.
 3. The diagnostic system of claim 1, wherein said diagnostic module monitors a percentage of sampled error over a defined period of time.
 4. The diagnostic system of claim 3, wherein said predetermined fault value is an accepted offset range.
 5. The diagnostic system of claim 4, wherein said diagnostic module determines an error value percentage indicative of said percentage of sampled error within said accepted offset range for said defined period of time.
 6. The diagnostic system of claim 5, wherein said diagnostic module indicates said electronic stepper regulator valve is stuck open when said error value percentage is approximately zero percent, said valve position of said regulator valve is approximately zero percent for approximately one hundred percent of said defined period of time, said evaporator coil inlet temperature is less than approximately 32° F., and a superheat value is approximately greater than 5° F.
 7. The diagnostic system of claim 6, wherein said superheat value is a moving average of a difference between said evaporator coil exit temperature and said evaporator coil inlet temperature to approximate said superheat value.
 8. The diagnostic system of claim 5, wherein said diagnostic module indicates said regulator valve is stuck closed when said error value percentage is approximately zero percent, said valve position of said regulator valve is approximately one hundred percent for approximately one hundred percent of said defined period of time, said evaporator coil inlet temperature is approximately greater than 32° F., and a superheat value is approximately greater than 5° F.
 9. The diagnostic system of claim 8, wherein said superheat value is a moving average of a difference between said evaporator coil exit temperature and said evaporator coil inlet temperature to approximate said superheat value.
 10. The diagnostic system of claim 5, wherein said diagnostic system indicates that air flow to the evaporator is blocked or evaporator fans are not operating properly when said valve position of said regulator valve is approximately one hundred percent for approximately one hundred percent of said defined period of time, said error value percentage is approximately zero, a superheat value is approximately less than 5° F., said evaporator coil inlet temperature is approximately less than 25° F., and said evaporator coil exit temperature value is less than said evaporator coil inlet temperature for greater than fifty percent of said defined period of time.
 11. The diagnostic system of claim 10, wherein said superheat value is a moving average of a difference between said evaporator coil exit temperature and said evaporator coil inlet temperature to approximate said superheat value.
 12. The diagnostic system of claim 1, wherein said regulator valve is an electronic stepper regulator valve and said control signal for adjusting a valve position of said electronic stepper regulator valve is a variable duty cycle control signal in which said duty cycle is a function of demand for cooling.
 13. The diagnostic system of claim 1, wherein said diagnostic module monitors and compares at least one of the following conditions: said valve position of said regulator valve; an error value percentage indicative of the percentage of sampled error within an accepted offset range for a defined period of time; a moving average of said valve position for a defined period of time; a steady state loading percentage set equal to said moving average of said valve position for a defined period of time when said error value percentage is less than fifty percent; a discharge cooling fluid temperature; said evaporator coil inlet temperature; said evaporator coil exit temperature; a moving average of a difference between said discharge cooling fluid temperature and said evaporator coil inlet temperature; a moving average of a difference between said evaporator coil exit temperature and said evaporator coil inlet temperature to approximate a superheat value; and a length of time said evaporator coil exit temperature is less than said evaporator coil inlet temperature during a predefined period of time.
 14. A method for diagnosing a cooling system, said steps comprising: producing a control signal for adjusting a valve position of a regulator valve as a function of demand for cooling; detecting an ambient temperature measurement, a coil inlet temperature measurement, and a coil exit temperature measurement; and comparing said valve position and at least one of said ambient temperature measurement, coil inlet temperature measurement, and coil exit temperature measurement with at least one predetermined fault value indicative of a fault condition.
 15. The method of claim 14 further comprising the step of issuing an alert signal when said at least one of said ambient temperature measurement, coil inlet temperature measurement, and coil exit temperature measurement bears a predetermined relationship to said fault value.
 16. The method of claim 14 wherein said step of comparing said valve position includes monitoring a percentage of sampled air over a defined period of time.
 17. The method of claim 16 further comprising the step of approximating a superheat value as a moving average of a difference between said evaporator coil exit temperature measurement and said evaporator coil inlet temperature measurement.
 18. The method of claim 17 further comprising the step of indicating said regulator valve is stuck open when said error value percentage is approximately zero percent, said valve position of said regulator valve is approximately zero percent for approximately one hundred percent of said defined period of time, said evaporator coil inlet temperature is less than approximately 32° F., and said superheat value is approximately greater than 5° F.
 19. The method of claim 16 further comprising the step of indicating said regulator valve is stuck open when said error value percentage is approximately zero percent, said valve position of said regulator valve is approximately one hundred percent for approximately one hundred percent of said defined period of time, said evaporator coil inlet temperature is approximately greater than 32° F., and said superheat value is approximately greater than 5° F.
 20. The method of claim 16 further comprising the step of indicating that air flow to the evaporator is blocked or evaporator fans are not operating properly when said valve position of said regulator valve is approximately one hundred percent for approximately one hundred percent of said defined period of time, said error value percentage is approximately zero, said superheat value is approximately less than 5° F., said evaporator coil inlet temperature is approximately less than 25° F., and said evaporator coil exit temperature value is less than said evaporator coil inlet temperature for greater than fifty percent of said defined period of time.
 21. The method of claim 14 wherein said step of adjusting a valve position includes providing a variable duty cycle control signal as a function of demand for cooling.
 22. The method of claim 14 wherein said step of monitoring and comparing a valve position includes monitoring and comparing at least one of the following conditions: said valve position of said regulator valve; an error value percentage indicative of the percentage of sampled error within an accepted offset range for a defined period of time; a moving average of said valve position for a defined period of time; a steady state loading percentage set equal to said moving average of said valve position for a defined period of time when said error value percentage is less than fifty percent; a discharge cooling fluid temperature; said evaporator coil inlet temperature; said evaporator coil exit temperature; a moving average of a difference between said discharge cooling fluid temperature and said evaporator coil inlet temperature; a moving average of a difference between said evaporator coil exit temperature and said evaporator coil inlet temperature to approximate a superheat value; and a length of time said evaporator coil exit temperature is less than said evaporator coil inlet temperature during a predefined period of time. 